The invention relates to the technical field of control valves and relates, according to its class, to a valve part of a control valve activated by an actuator for controlling flows of pressurized medium.
In internal combustion engines, gas-exchange valves are activated by the cams of a camshaft set in rotation by the crankshaft, wherein, through the arrangement and shape of the cams, the control times of the gas-exchange valves can be set in a selective way. From the background of thermodynamic processes, it has proven advantageous if, during the operation of the internal combustion engine, the control times of the gas-exchange valves are influenced as a function of the current operating state of the engine, such as rotational speed or load. This influence is set by the relative rotational position between the cam and crankshaft. The use of devices for changing and fixing the relative rotational position between the camshaft and crankshaft, generally designated as “camshaft adjusters,” has been known for a long time.
Camshaft adjusters typically comprise a drive part locked in rotation with the crankshaft via a drive wheel and a driven part fixed to the camshaft, as well as a hydraulic control drive, which is connected between the drive part and driven part and which transmits the torque from the drive part to the driven part and allows a fixing and also adjustment of the relative rotational position between the drive part and driven part.
Hydraulic camshaft adjusters are typically constructed as axial piston adjusters or rotary piston adjusters. For an axial piston adjuster, the drive part engages with a piston via helical gearing. This piston engages, on its side, with the driven part by helical gearing. Between the drive part and driven part, a pressure space is formed, which is divided by the piston into two pressure chambers. For a rotary piston adjuster, the drive part constructed in the form of an external rotor (“stator”) and the driven part constructed in the form of an internal rotor (“rotor”) are arranged concentrically and adjustable in rotation relative to each other. Pressure spaces are formed in the radial intermediate space between the stator and rotor. A vane connected to the rotor extends into each of these pressure spaces, such that each pressure space is divided into two pressure chambers. Through selective pressurization of the pressure chambers of each pressure space, that is, by generating a pressure difference across the pressure chamber pair of each pressure space, the drive part can be moved relative to the driven part, so that a rotation of the camshaft and consequently a change in the relative rotational position between the camshaft and crankshaft is created. On the other hand, the relative rotational position can be maintained through a corresponding equal pressurization of the two pressure chambers of a pressure space.
Controlling the hydraulic camshaft adjuster is realized by a control unit, which controls the feed and discharge of pressurized medium to and from the individual pressure chambers based on detected characteristics of the internal combustion engine. The flows of pressurized medium are regulated by a control valve (proportional valve) controlled by the control unit.
Control valves for controlling the flows of pressurized medium for camshaft adjusters have been known as such for a long time and are described, for example, in the European Patent Application EP 1 596 041 A1 and the German Offenlegungsschrift [unexamined patent application] DE 102 39 207 A1 of the applicant. They comprise, as essential components, an actuator, typically an electromagnet with a hollow cylindrical magnetic housing, in whose hollow space a coil winding and an axially moving magnetic armature are arranged with a tappet, as well as a hydraulic valve part with a hollow cylindrical valve housing, in whose hollow space a control piston that can move in the axial direction is held. When the magnetic armature is energized, the tappet acts on the control piston of the valve part, so that the piston can be displaced in the axial direction against the compressive force of a compression spring, in order to regulate the flows of pressurized medium in this manner.
In a typical construction, the valve housing is provided on its outer periphery with a plurality of annular grooves that are spaced apart in the axial direction and in which radial boreholes are machined. These boreholes open into the hollow space of the valve housing and are used as a pressure connection and work connections. The control piston can be provided in the form of a hollow piston with a hollow space, which is open on one side and whose opening is used as a discharge connection. If the hollow space opening of the control piston is located on the end away from the tappet, then it can be formed as an axial opening. If the hollow space opening of the control piston is located on the end facing the tappet, then it is necessary to form this opening as a radial opening, in order to provide a sufficient contact surface on the control piston for the tappet. An example construction of such a valve part is shown in FIG. 4.
Accordingly, the valve part designated as a whole with the reference number 100 of an electromagnetic control valve comprises a hollow cylindrical valve housing 101, which surrounds a valve housing hollow space 103 with an axial hollow space opening 121. In the valve housing hollow space 103, a control piston 102 is held so that it can move in the axial direction. A tappet 104, which is only shown partially and which is attached rigidly to a magnetic armature of an electromagnet not shown in FIG. 4, contacts the end face 105 of the control piston 102 at the left in FIG. 4. When the magnetic armature is energized, the tappet is displaced in the axial direction relative to the valve part 100 and in this way displaces the control piston 102 against the spring force of a compression spring 106. On one end, the compression spring 106 contacts the end of the control piston 102 away from the tappet and for this purpose is held in an axial first ring step 107. On its other end, the compression spring 106 is supported on a base surface 109 oriented perpendicular to the axial direction of an axial second ring step 108 of the valve housing hollow space 103.
The valve housing 101 is provided on its outer periphery with three ring grooves, namely a first ring groove 124, a second ring groove 125, and a third ring groove 126, spaced apart in the axial direction. In the ring grooves, first radial boreholes 110, second radial boreholes 111, and third radial boreholes 112 are machined uniformly about the periphery, which each open into the valve housing hollow space 103. In the shown axial section, the ring grooves transition directly into the radial boreholes, so that they are not distinguished from the ring grooves in the drawing. As indicated by the arrows, the first ring groove 124 with the first radial boreholes 110 acts as a first work connection A, the second ring groove 125 with the second radial boreholes 111 acts as a pressure connection P, and the third ring groove 126 with the third radial boreholes 112 as a second work connection B.
The control piston 102 is constructed in the form of a hollow piston, wherein the control piston hollow space 118 is formed by a blind borehole open toward the end face 109 of the valve housing. In the outer periphery of the control piston 102, three ring grooves are machined, namely a fourth ring groove 114, a fifth ring groove 115, and a sixth ring groove 113 located between the fourth and fifth ring groove. The fourth ring groove 114 is provided with fourth radial boreholes 116 distributed uniformly about the periphery and the fifth ring groove 115 is provided with fifth radial boreholes 117, which are distributed uniformly about the periphery and which each open into the control piston hollow space 118. Furthermore, the control piston 102 is provided on its tappet-side end section with sixth radial boreholes 119, which are arranged distributed about the periphery and which connect the control piston hollow space 118 with a twice offset, axial third ring step 120 in a fluid-conducting way, which opens into the hollow space opening 121 of the valve housing 101. The hollow space opening 121 is used as a discharge connection T. Adjacent to the sixth ring groove 113 are a first ring bar 122 and a second ring bar 123, whose peripheral surfaces are shaped so that for an axial displacement of the control piston 102, the first and the third radial boreholes 110, 112 can be covered and opened, in order to regulate the flow rate of pressurized medium in this way by changing the cross sections of the openings.
Thus, according to the axial position of the control piston 102, the first work connection A and the second work connection B can be connected in a fluid-conducting way selectively with the pressure connection P or the tank connection T. In FIG. 4, a situation is shown, in which the first work connection A is connected to the tank connection T, while the second work connection B is connected to the pressure connection P. If pressurized medium flow into the control piston hollow space 118 (as specified by the dashed arrows), the flows of pressurized medium directed inward in the radial direction are deflected into an axial flow of pressurized medium, which is directed toward the tappet-side end of the control piston hollow space 118 and which flows essentially in the middle of the control piston hollow space 118. Then the axial flow of pressurized medium is deflected into flows of pressurized medium directed outwardly in the radial direction, which are diverted into the discharge connection T. In this way, inevitably a build-up pressure is generated by the deflection of the axial flow of pressurized medium on the end face of the control piston hollow space 118, which applies a load on the control piston 102 in a direction, which is equal to the direction of the spring force of the compression spring (to the left in FIG. 4). In addition, an eddy flow is generated by the deflection of the axial flow of pressurized medium. Consequently, these effects lead to an imbalance in the pressure forces primarily in the middle of the control piston 102, with these forces generating the undesired force curves for the desired axial displacements of the control piston 102. In addition, the tappet 104 activated by the electromagnet must shift the control piston 102 against a higher resistance, so that the electromagnet must have a sufficiently robust construction, in order to withstand an increased heat generation for the higher current intensities necessary for this purpose.